Magnetic bearings are subject to performance limits which are quite different from those of conventional bearings. These are due in part to the inherent nonlinearity of the device and in part to its electrical nature. Three important nonideal behaviors are presented: peak force capacity, force slew rate limitation, and sensitivity to rotor motion at large displacements. The problem of identifying the dynamic requirements of a magnetic bearing when used to support a known structure subject to known loads is discussed in the context of these limits. Several simple design tools result from this investigation.

Magnetic journal bearings are coming into increasing use in industry today. They are primarily used to replace either rolling element or fluid film bearings in rotating machinery. The major advantages are elimination of oil systems and associated seals, expected very long life, very low power losses, and great potential for vibration reduction. Disadvantages include lack of field experience, unknown reliability over a long time, high cost (so far primarily due to the small quantity being made for a given application), and advanced automatic control design required. This paper discusses the design of a magnetic journal bearing with four electromagnets arranged radially around a shaft to fully support a rotor. Each electromagnetic is connected to a controlling electronics circuit which regulates the current to the magnet. For the measurements presented here, only the top magnet was tested and the shaft was not rotating. Thus a single control algorithm was isolated from other effects. This paper compares two control algorithms with differing circuit band widths of 1.2 kHz and 50 kHz. The wider bandwidth algorithm produced approximately a sixfold increase in magnetic bearing stiffness and a much greater stable operating region compared to the lower bandwidth algorithm. Overall, the calculated effective stiffness and damping coefficients were within 20 to 30 percent of the measured values.

No abstract

No abstract

Many r o t a t i n g machines such as compressors, t u r b i n e s and pumps have l o n t h i n s h a f t s with r e s u l t i n g v i b r a t i o n problems. They would b e n e f i t from a d d i t i o n a f damping n e a r t h e c e n t e r of t h e s h a f t . t h e s e machines because t h e y can o p e r a t e i n t h e working f l u i d environment u n l i k e conventional b e a r i n g s . This paper d e s c r i b e s an experimental test r i g which was set up with a long t h i n s h a f t and s e v e r a l masses t o r e p r e s e n t a f l e x i b l e s h a f t machine.An active m a n e t i c damper was placed i n t h r e e l o c a t i o n s : n e a r t h e midspan, n e a r one midspan l o c a t i o n reduced t h e f i r s t mode v i b r a t i o n 82X, t h e d i s k l o c a t i o n reduced it75% and t h e b e a r i n g l o c a t i o n a t t a i n e d a 74% reduction. Magnetic damper s t i f f n e s s and damping v a l u e s used t o o b t a i n t h e s e reductions were only a few percent of t h e bearing s t i f f n e s s and damping v a l u e s . A t h e o r e t i c a l model of both t h e r o t o r and t h e damper was developed and compared t o t h e measured r e s u l t s . Magnetic dampers have t h e p o t e n t i a l t o be employed i n end d i s k , an 8 c l o s e t o t h e bearing. With t y p i c a l c o n t r o l parameter s e t t i n g s , t h eThe agreement was good.